Variable-displacement turbine-speed hydrostatic pump



Oct. 29, 1968 Filed May 12 RR-MITCHELL ETAL VARIABLE-DISPLACEMENT TURBINE-SPEED H1 'DROSTATIC PUMP 5 Sheets-Sheet l JAMES C. SWAIN "JOHN P; WILCOX mvsu'rons nrronusvs' I ROBERT K-IMfl CHELL- 6 Sheets-Sheet'Z R. K. MITCHELL ET AL Oct. 29, 1968 VARIABLE-DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed May 12, 1966 INVENTORS svfimy zaz dmaw), ATTORNEYS Oct. 29, 1968 MITCHELL ET AL VARIABLE-DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Filed May 12, 1966 STRAIGHT LINE MOVABLE MEMBER IQS v v m FULL DISPLACEMENT Ish/ ACTUATOR MOTION Foa- FULL msPLAcEMEm 6 Sheets-Sheet ROBERT K. MITCHELL JAMES C. SWAIN DAVID L. THOMAS JOHN P. WILCOX nvvsmrons v v BYj/ f,

0M ATTORNEYS 2 9, 1968 I R. K. MITCHELL; ETAL v $307,742";

VARIABLE-DISPLACEMENT TURBINE-SPEED HYDROSTAT'IC PUMP Filed May 12, 1966 4 Y H a sheets-sheet 4" [JAMES c. swm

DAVID I. moms JOHN P WILCOX l wvznrons j bhufjfhamub I "Wm;

ROBERT mm;

Oct. 29, 1968 M|THELL ET AL 3,407,742

VARIABLE-DISPLACEMENT TURBINE-:SPEED I'IYDROSTATIG PUMP Filed May 12, 1966 6 Sheets-Sheet 6 ROBERT K MITCHELL JAMES c. SWAIN JOHN F? WILCOX 'INVENTORS Fiq.*9b DAVID L. THOMAS ,OWATTORNEYS United States Patent 3,407,742 VARIABLE-DISPLACEMENT TURBINE-SPEED HYDROSTATIC PUMP Robert K. Mitchell, James C. Swain, David L. Thomas, and John P. Wilcox, Columbus, Ohio, assignors to The Battelle Development Corp., Columbus, Ohio, a corporation of Delaware Filed May 12, 1966, Ser. No. 549,679 8 Claims. (Cl. 103-120) ABSTRACT OF THE DISCLOSURE A sliding-vane rotary pump operating at high rotational speeds and constructed for variable displacement. The pump housing includes two lobes that are enlargeable or reducible to provide variable displacement while maintaining all portions of the cam surface (that surface along which the vane tips track very closely) in tangency. Moveable lap spaces are connected by a plurality of bridges across the pump ports to nonmoveable lap spaces. The ports are positioned along the cam surface rather than in the end plates and the ratio of rotor length to rotor diameter is considerably greater than in conventional sliding-vane rotary pumps. The van tips are pivotal and supported by a hydrodynamic film of oil between the vane tip and vane-tip track which is made possible by the high rotational speeds at normal operating conditions. The tip, vane, and vane socket in the rotor are constructed to react to and control various forces acting on the rotating vanes such as centrifugal force, pressures in front of and behind the vane, socket pressures, bending moments, tipping forces, hydrostatic pressures, etc.

This invention relates to a variable-displacement sliding-vane pump for operation at high rotational speeds. More particularly, it concerns a sliding-vane rotary pump operable at high turbine speeds and the invention is especially useful as an element of a transmission for a turbinepowered vehicle.

One of the difficult problems associated with adopting a high-power turbine to drive a wheeled vehicle is the efficient utilization and control of the high rotational turbine-shaft speeds that are usually in excess of 20,000 rpm. The speed of a high-power turbine is relatively unresponsive to load and fuel changes because of the large inertia of the rotor. This speed is much higher than can be used directly at the wheels; therefore, a

speed reducing or power conversion mechanism is necessary. For a responsive vehicle, which requires a wide range of speeds at constant power, the speed reducing mechanism must be an elficient variable speed transmission. A hydrostatic transmission, having variable displacement and infinite variability across its range, is a desirable type of transmission for a constant-speed turbine. This invention provides a pump that supplies these desired features.

Although the pump described herein has other applications, it will be discussed mainly with respect to its use with a turbine and the problems associated therewith.

Numerous sliding-vane rotary pumps have been suggested in the past. The sliding-vane rotary pump includes a rotor surrounded by a housing that may be circularshaped and eccentric to the rotor, or the housing may be elliptical in shape having the two ends of the ellipse serving as lobular pumping chambers. The rotor includes a plurality of slots or sockets with a vane (sometimes called a piston) slidably mounted in each slot. The tip of the vane follows the inner contour of the housing primarily due to the acceleration forces generated by rotation of the rotor. Pumping action occurs within two vanes (or a vane and a point of close proximity between the rotor "ice and the housing), the housing, rotor and the end plates on the housing. The inlet of the pump is generally positioned to supply fluid at a point where the vanes move radially outward and the outlet is positioned at a point where the vanes move radially inward.

There are also a number of sliding vane rotary pumps that have been constructed to have variable displacement. Two methods of varying the displacement appear most frequenlty. In one method, the eccentricity of a circular housing and the rotor are varied with respect to one another. The other method involves varying the size or shape of the pump housing by various means. None of these prior pumps are practical for high-speed operation.

The apparatus of this invention provides a two-lobed sliding-vane rotary pump wherein the lobe spaces are enlargeable or reducible to provide variable displacement. Preferably, the construction of the cam surface or vanetip track includes two fixed surfaces (lap spaces), two movable surfaces (lap spaces) and a plurality of connecting links or bridges. An important feature that distinguishes the construction of the pump of this invention from prior pumps that appear to be similar in construction is that the cam surfaces of the bridges are maintained in tangency (at the intersection points), throughout all displacement changes, to the surfaces of the fixed and movable lap space where the bridges interconnect them. The passage of vanes from the fixed and movable surfaces to the bridge surfaces, and vice versa, is always accomplished smoothly regardless of the displacement settings. The feature of keeping the vanetip track elements tangent at all pump displacements is very important as the high speed exaggerates the slightest bump or discontinuity in the vane-tip track causing skips, bounces, leakage, high wear and early failure of the pump.

Most conventional two-lobed pumps have the ports located in the end plates of the pump housing. In this invention, the ports are preferably positioned so that the flow enters and leaves the pump through spaces between the interconnecting bridges along the complete length of the rotor. This port construction, through the bridges, has the effect of minimizing flow losses through the pump. Preferably, the ratio of porting space to vane support (bridge surface) is about sixty percent open area and about forty percent bridge area. The selection of this ratio is a compromise between the need to provide a suflicient vane-tip hydrodynamic bearing area and the desire to minimize flow losses.

This invention provides a sliding-vane rotary pump that is two-lobed, hence pressure balanced. The flow rate through the pump is relatively free from pulsations and the pump has a very small clearance volume at the hlghpressure, low-flow condition which is important in minimizing compressibility effects at high pressure.

This invention further provides a sliding-vane rotary pump wherein the ratio of rotor length to rotor diameter is considerably greater than in conventional sliding-vane rotary pumps. The small-diameter, long rotor construction fits in very well with the variable-displacement pressurebalance'd construction since the bearing loads and rotor distortions are minimized.

One advantage of the small-diameter, long-axis rotor construction for the pump of this invention is the potentially high overall efficiency. The most significant loss in the pump is the flow pressure drop loss which is essentially a square function of vane tip velocity. At any selected angular speed, the vane tip velocity, and therefore pump flow, is a direct function of diameter. Therefore, because flow through the pump is in the turbulent area, flow losses are essentially a square function of rotor diameter. Vane tip friction loss is a function of vane tip speed and vane tip load. Both vane tip speed land vane tip load are directly related to the diameter at a given angular speed so that vane tip friction loss is also a square function of the diameter. Other losses, such 'as hearing loss and viscous drag on the ends of the rotor, are also reduced by small rotor diameter.

Another advantage of the invention is that the smalldiameter, long-rotor of the pump minimizes charging pressure. A charging pressure is required to accelerate the fluid being pumped to vane tip velocity without cavitation occurring. This charging pressure is a square function of vane tip velocity so that a rotor of twice the diameter of another rotor rotating at the same angular speed requires four times the charging pressure. Thus, the smaller diameter rotor has a distinct advantage.

In most conventional sliding-vane rotary pumps the vane tip contacts the cam surface or vane track in a substantially sliding relationship. The lubrication is in the nature of boundary lubrication. A number of pivoting vane tips have been used in prior vane pumps essentially to aid in sealing, to provide a larger load carrying surface, or to distribute the wear on the tip more evenly. Usually the conventional vane tip is constructed with an arc radius substantially equal to the radius of the vane track surface.

In this invention, the vane tip does not touch the vane track but rides on a wedge of oil with friction being about ten times less than that of boundary lubrication This is possible because the pump of this invention is operated as a high-speed pump providing vane tip velocities sufiicient to form a thin but very stiff hydrodynamic film of oil between the vane tip and the vane-tip track. The vane tip is pivotally mounted in the end of the vane and the contact radius of curvature of the vane tip is selected so as to be smaller than the smallest radius existing in the vane-tip track. There are other unique construction features of the vane and pivoting vane-tip that react to and control various other forces acting on the rotating vanes; for example, such forces as centrifugal force, pressures in front of and behind the vane, socket pressures, bending moments, tipping rforces, hydrostatic pressures, etc.

An object of this invention is to provide a sliding- Wane rotary pump that is capable of variable displacement and that is relatively simple in construction.

Another object of this invention is to provide a variable-displacement pump that has a smooth, continuous displacement variation.

Another object of this invention is to provide a pump with both hydraulic pressure and dynamic balance to reduce bearing loads and distortion of rotating parts.

Another object of this invention is to provide a variable-displacement, sliding vane rotary pump with reduced motion of internal parts and reduced leakage area at reduced displacement.

Still another object of this invention is to provide a variable-displacement, sliding-vane rotary pump especially capable of operating at high-rotational speeds.

Still another object of this invention is to provide variable-displacement, sliding-vane rotary pump with smooth flow paths to minimize losses.

Still another object of this invention is to provide a high-capacity, variable-displacement, sliding-vane rotary pump with a high length-to-diameter ratio that minimizes losses and improves efiiciency.

Still another object of this invention is to provide a vane-tip construction that adapts to being supported by a hydrodynamic oil film.

Still another object of this invention is to provide a vane construction that minimizes the various loads imposed on the vane and the vane tips.

Still other objects and advantages of this invention will be apparent from the description that follows, the drawings and the appended claims.

In the drawings:

FIG. 1 is a sectional elevational view of the pump -of this invention taken along the line 11 of FIG. 2;

FIG. 2 is a sectional elevational view of the pump of this invention taken along the line 22 of FIG. 1;

FIGS. 3a and 3b are diagrams illustrating the relationship of the various cam track elements;

FIG. 4 is a sectional view of the bridges and a port;

FIG. 5 is a perspective view of two of the bridges and a portion of an associated fixed cam track member;

FIG. 6 is a sectional view of the apparatus for ensuring equalization of the displacement of the two lobes of the pump chamber;

FIG. 7 is an enlarged sectional view of a portion of the rotor, a vane and a portion of the cam surface;

FIGS. 8a and 8b are diagrams (enlarged) of a vane tip showing the various forces that act thereon; and

FIGS. 9a and 9b are diagrams (enlarged) of a vane showing the various forces that act thereon.

Referring to FIGS. 1 and 2, the pump 21 has a pump housing 23 and a manifold section 25 attached at one end. A housing flange 27 and a manifold flange 29 are held together by suitable means such as bolts 3131. The contacting surfaces of pump housing 23 and a manifold section 25 are provided with sealing means; preferably Orings 3333, that seal around the fixed end plate 37 and also around the outlet passages 3939 and inlet passages 41-41 that communicate between housing 23 and manifold 25. The fixed end plate is an integral part of the manifold section 23 and closes off one end of the variable displacement pump chamber 42.

The opposite end of the pump housing 23 is provided with a retainer 43 held onto the housing 23 by a plurality of bolts 45-45. The retainer 43 has an outwardly extending portion or pump mounting flange 47. A hearing housing 49 fits inside a central opening 50 of the retainer 43 and is attached to the retainer 43 by bolts 51-51. An O-ring 53 seals between the retainer 43 and pump housing 23 and an O-ring 55 seals between the retainer 43 and bearing housing 49.

Pumping energy is transmitted through an input shaft 57, to the rotor 59 and then to each individual vane 61. The input shaft 59 has a spline 63 to receive energy from a power source (such as a turbine, not shown) and a spline 65 engages the rotor 59 at about the midpoint section along the longitudinal axis of rotor 59. The preferred attachment arrangement aids in reducing torsional stresses and deflection in the rotor 59 that would occur if the shaft rotational force were applied to either end of the rotor 59. Such a condition could also exist if the input shaft 57 were arranged for engagement along the length of rotor 59 and only one point or section of the engaging surfaces actually transferred the rotative energy. The rotor 59 is radially positioned and supported by a bearing 67 positioned in the fixed end plate 37 and a sleeve and thrust bearing 69 positioned in a pressure-loaded end plate 71. The pressure-loaded end plate 71 and fixed end plate 37 axially position the rotor 59. A collar 73 is threadedly engaged with the input shaft 57 and serves to position and clamp the rotating member 75 of a mechanical face seal 77. The rotating member 75 is clamped against an annular shoulder 79 on the shaft 57. An annular projection 81 on the collar 73 also positions the input shaft 57 axially against sleeve and thrust hearing 83 so that the face seal 77 functions. The collar 73 receives the thrust due to the pump charging pressure which tends to force the input shaft 57 outward. The input shaft 57 is positioned radially by the sleeve and thrust bearing 83, for which the collar 73 is the journal, and by the rotor 59. The fixed member 85 of the face seal 77 is attached to the bearing housing 49. An O-ring seal 87 is provided between rotating face seal member 75 and the input shaft 57, and suitable retaining and seal means 89 is provided for sealing between fixed member 85 and the retainer 49. The mechanical face seal 77 seals the pump charging pressure.

As the rotor 59 turns, the vanes 61 are guided and caused to stroke by the cam surface 90. The cam surface 90 is comprised of lap spaces 91 and 93 of fixed members 95 and 97, the surfaces 99-99 of a plurality of independent bridges 101-101 and lap spaces 103 and 105 of movable members 107 and 109 (FIGS. 2, 3a, 3b, and 4). As seen in FIG. 4 the bridges 101-101 are not closely stacked, but are preferably separated leaving about fifty or sixty percent of the area between the lap spaces 91, 93, 103, and 105 open. It is through these openings that the fluid enters and leaves the pump chamber 42. A plurality of flow passage relief spaces 111-111 are provided on each of the members 95 and 97 (FIGS. 2 and 4). These provide the last exits and entrances for pressurized fluid to exit from in front of a vane 61 on the outlet side and for fluid to first enter behind a vane 61 on the inlet side.

The movable members 107 and 109 are movable in a radial direction, with respect to the rotor 59 to vary the displacement or size of pump chamber 42. The bridges 101-101 are positioned by tangs 113-113 on members 95, 97, 107, and 109 fitting closely into slots 115-115 in the bridges 101-101. The bridge 101-101 are positioned axially by their ends 120-120 fitting closely into slots 117-117 on the members 95, 97, 107, and 109. Since each bridge 101 is subjected to the operation of two nonparallel tangs 113-113, its position is uniquely determined for each displacement setting. Portions of the lap space parts 91, 93, 103, and 105 overlap the sides of the bridges 101-101 and serve to carry the vanes 61-61 across the gaps in the cam surface 90 which are formed as the bridges 101-101 slide away from one lap space or another. The cam surfaces on these extended portions 119-119 of the lap space (95, 97, 105, and 107) are planar so that they appear as straight lines (FIGS. 3a and 3b). Since these lines (119-119) are parallel to the tangs 113-113, the corners 121-121 of the bridges 101-101 are constrained to be on the lines (119-119). And since in addition, the cam surfaces 99-99 of the bridges 101-101 are constructed to be tangent to the lines (119-119) at the corner points 121-121, the passage of the vanes 61-61 from lap space to bridge and from bridge to la-p space is always made smoothly, irrespective of the displacement settings.

The double-tang construction for bridge positioning has distinct advantages over other methods wherein the bridges are pivoted with respect to one or both of the lap spaces. In other constructions the rotation of the bridges with respect to the lap spaces leads to unfavorable conditions as the vanes pass from one to the other. If sharp corners (against which the vanes would impact) are to be avoided in such construction, when the displacement setting is at one extreme, then a reverse curvature in the cam surface exists when the displacement setting is at the opposite extreme. The vanes come out of contact with the cam surface at this point and must be loaded to prevent such skips by some means other than their own centrifugal force. This loading is not only difficult to achieve, but it also increases the maximum tip pressures imposed on the vanes as they cross adjacent areas of the cam surface that have high positive curvature.

FIGS. 3a and 3b show the relationship of a fixed member 95, a bridge 101 and a movable member 109 at full and zero displacements. The position of the pump elements in FIG. 3a is designated as zero displacement" because with this position of the cam track sections, there is no output flow from the pump. The radii of the curved cam track sections are different and, as shown in FIG.

311, there are straight sections 119-1*19 that are a portion position in the rotor 59 to some degree even at zero displacement. In the embodiment shown in FIG. 6a, the small stroking movement occurs as the vanes 61 cross the [bridges 101-101 and adjacent straight portions 119-119. In FIGS. 3a and 3b the radii of the rotor 59 and lap spaces (91, 93, -103, and 105) are substantially equal. Their axes coincide along a central axis 123 (shown as a point in FIGS. 3a and 3b) when the pump is at zero displacement. However, other arrangements are also workable and it is sometimes preferred to select unequal radii as discussed subsequently. The radii of the bridge surfaces 99-99 is preferably smaller than the radii of the rotor 59 and lap spaces 91, 93, 103, and 105. The axes 125-125 of the bridge surfaces 99-99 are posi tioned at about equidistant points around the central axis 123 when the pump is at zero displacement (one of which is represented as a point 125 in FIGS. 3a and 3b). A line 127, perpendicular to the straight section 119 at the point where lap space 91 changes from curved to straight section 119, passes through the central axis 123. Another line 129, perpendicular to straight section 119 at the point where lap space 105 changes from curved to straight section 119, also passes through the central axis 123. Lines 131 and 133, perpendicular to straight sections 119-119 at the corner points 121-121 of the bridge 101, intersect at the axis 125. The above intersections indicate tangency of all elements of the cam track with the straight sections 119-119 being tangent to the curved elements 91, 99, and 105, and this, of course, is similarly true for the other elements of the cam track 90 that are not shown in FIGS. 3a and 3b.

In FIG. 3b, the movable member 109 has moved the lap space outwardly (to full displacement) as indicated by the space between lap space axis 135 and central axis 123. The perpendicular line 127 still intersects central axis 123 and perpendicular line 131 still intersects axis 125. Perpendicular lines 129 and 133 coincide and intersect both axis and axis 135. As lap space 105 moves outwardly, bridge 101 moves away from fixed member 95, but moves toward movable member 109. This is due to the tang and slot (113 and 115) arrange ments which are positioned with their sides parallel to the straight sections 119-119. The slot 115 hearing against the side of tang 113 on fixed member 95 forces bridge 101 toward movable member 109 as movable member 109 moves outwardly. Thus, regardless of the position of movable members 107 and 109 the surfaces (91, 93, 99-99, 103, and 105) of the cam track 90 are maintained in tangency.

The radii of lap spaces 91, 93, 103, and 105 are shown to be equal (FIGS. 3a and 3b) as one example of the construction. The radii are selected depending on the pump application. It is desirable that lap spaces 91 and 93 have their center of radius at point 123 so that no stroking of the vanes 61-61 relative to the rotor 59 occurs while the vanes are traversing the field or immovable lap spaces 91 and 93. In FIG. 317 it is evident that stroking of the vanes 61-61 would occur in the lap space 105 because the center of the lap space are 135 and the center of the rotor 123 do not coincide. The radius of the lap space 105 and 107 should be selected :to minimize vane stroking in these lap spaces depending upon the specific application of the pump. In some applications the axes of the movable lap spaces 105 and 107 will not coincide at the axis 123 at zero displacement (the radii of these lap spaces may be selected to be larger) anrd stroking of the vanes 61-61 will occur as they cross the movable lap spaces 105107.

FIG. 2 shows the construction for moving the lap spaces 103 and 105 toward and away from the central axis 123. The apparatus that moves the member 107 is duplicated for moving member 109 so that the apparatus will be described only with respect to its relationship to member 107. The movable member 107 is attached to a pressure balance piston 137 by suitable attachment means such. as bolts 139-139. The balance piston 137 includes a seal slot 141 partially enclosed by a base section 143. The balance piston 137 is reciprocal in a chamber 145 that is closed off by a cylinder section 147. The cylinder section 147 is attached to the pump housing 23 by bolts 149-149. The seal slot 141 is provided with a seal 151 that seals between the balance piston 137 and the outer wall of the chamber 145. Another seal 153 is positioned to seal between the balance piston 137 and the movable element 107. The balance piston 137 is provided to essentially balance the pressure load on movable member 107 resulting from pressure generated in the pump chamber 42 and to reduce the force required to move the memher 107 toward zero displacement against the pump pressure. High pressure fluid in the discharge port 155 is communicated to the chamber 145 behind the balance piston 137 through passage 157 and ports 159159. Thus the force (due to pump pressure) that tends to force the movable member 107 away from the rotor 59 is substantially neutralized. The lap space member 107 must be in intimate contact with the case 23 to minimize leakage from the high pressure in the discharge port 155. Position of the sealing land 160, on movable member 107, is selected so pressure forces always cause intimate contact and some nominal load between the sealing lands and the pump housing 23. The axis of the pressure balance piston 137 must also be positioned so that pressure forces always cause intimate contact and some nominal load between the sealing land 160 and the case 23. In positioning both the sealing land 160 and the axis of pressure balance piston 137 it must be recognized that the area over which pressure acts on lap space surface 103 varies with time because of movement of any vane 61 across the lap space which acts as a sealing member (i.e., the area of pressure application on the lap space surface is varied as the seal ing vane 61 moves).

The balance piston 137 is attached to a rod 161 that passes through the wall 163 of cylinder section 147 into cylinder 165. Appropriate seals 167167 are provided around rod 161 to seal cylinder 165 from balance chamber 145. A double acting piston 169 is positioned in cylinder 165 and is attached to rod 161 by a nut 171 that forces piston 165 against a shoulder 173 on the rod 161. An O-ring 175 is provided around piston 169 to seal against the wall of cylinder 165. The cylinder 165 is closed off by a cylinder head 177 also provided with an O-ring seal 179. Ports 181 and 183 are provided at each end of the cylinder 165 to admit fluid under pressure thus actuating the piston 169 and moving member 107. Fluid admitted through port 181 moves member 107 toward rotor 59 and zero displacement and fluid admitted through port 183 increases th pump displacement.

In FIG. 1, the pressure loaded end plate 71 is radially positioned by the housing 23. The end plate 71 has an annular-shaped extension or annular piston 185 that fits into an annular cavity 187 positioned between the re tainer 43 and bearing housing 49, with O-ring seals 189-- 189 provided on the piston to seal off the open side of the cavity 187. At least one port 191 passes through the annular piston 185 and delivers the outlet pressure of the pump to chamber 187 thus pressure-loading the plate 71 and uring it toward the rotor 59 and against the ends of fixed lap spaces 95 and 97 of movable lap spaces 107 and 109, and against shoulders 192 and 194 on the case 23. The end plate 71 is pressure loaded to maintain a small nominal rotor end clearance and to act as a seal to minimize leakage from the discharge pressure passages 39 39. Though there must be intimate contact between the end plate 71 and the fixed lap spaces 95 and 97, the movable lap spaces 107 and 109 and the case at shoulders 192 and 194, the loading must be such as to allow stroking of the movable lap space members with relative ease.

To keep pressure forces acting on the rotor 59 balanced, it is desirable to ensure that movable members 107 and 109 will always be substantially equidistant from the central axis 123 (or rotor 59). By maintaining the movable members 107 and 109 equidistant from central axis 123, the displacement of the two lobes of pump chamber 42 have equal displacement and equal flow losses resulting in pressure balance of the rotor. The construction for maintaining the movable members 107 and 109 equidistant from the central axis 123 at all displacements is partially shown in FIG. 1 and more completely in FIG. 6. The equalizer plate 193 is positioned between the pressureloaded end plate 71 and the retainer 43 for longitudinal positioning. The equalizer plate 193 is provided with two slots 197-197 with a slider 199 fitted into each slot 197. The slider is provided with a pin 201, and each pin passes through the end plate 71 and engages one of the movable members 107 or 109. The equalizer plate 193 is free to rotate on the bearing surface 195 on end plate 71 so that if one of the movable members, for example 107, moves toward the central axis 123, that movement is transferred through one of the pins 201 to equalizer plate 193 which causes the plate 193 to rotate and of necessity, through engagement of the other movable member 109 with the other pin 201, causes member 109 to move toward the central axis 123 an equal amount (the actual engagement of pins 201201 with movable members 107 and 109 is not shown).

The cam surface includes eight circular arcs (91, 93, 103, 105, and four 9999) connected by straight tangent surfaces (119-119) of variable length. Some of the operating problems of an example pump are: (l) rotational speeds of about 22,000 r.p.m.; (2) a pressure range of about 2000 to 8000 p.s.i.; (3) absorption of constant power over the range of pressures; (4) some stroking of the vane while the vane is sealing; (5) the vanes must cross over the ports in the bridges; (6) disappearance of centrifugal force and reduction of acceleration forces as the vanes cross the straight sections of the cam surface; and (7) full pump pressure differential across the vane while the vane crosses a lap space. The foregoing conditions require that the vanes 61-61 be of a more specialized construction than the vanes of conventional pumps. The relatively large hydrostatic pressures tend to force the vanes 61-61 away from the cam surface 90 and must be balanced. The vane tip is hydrodynamically lubricated, i.e., it is separated from the cam surface 90 by a thin film of oil.

The vane 61 and vane tip 203 mounted in the rotor 59 are shown in the enlarged cross-sectional view of FIG. 7. The vane tip 203 is a type of pivoted slider bearing mounted in a socket 205 of the vane 61. The slider surface 207 is relatively small and this accomplishes two things: (1) the hydrostatic force on the tip 203 is relatively small compared to the hydrodynamic force, there by allowing the hydrodynamic force to control the angle of tilt; and (2) hydrostatic radial vane forces can be balanced by an undervane step 209. The stepped vane 61 is mounted in a matching socket 211 so that a first chamber 213 is formed under the step 209 and second chamber 215 is formed under the vane end 217. The vane 61 moves in the direction of the arrow 219 so that the step 209 is considered to be on the front of the vane 61. The pressure in chamber 42a, in front of the vane 61, communicates with the chamber 213 by means of a port 221 and passage 223. The pressure in chamber 42b, in back of the vane 61, communicates with chamber 215 by means of a port 225 and passage 227. The undervane area of step 209 is larger than the undervane area of the vane end 217 because the hydrostatic force on the slider surface 207 (represented by force 249 in FIG. 8b) is greater when outlet pressure occurs in pumping chamber 42a than when outlet pressure occurs in pumping chamber 42b due to the forward tilt of the tip 203. The vane 61 is just wide enough to provide suflicient strength in the tip socket 205 to hold the tip 203 against hydrostatic forces. The moment due to tip hydrodynamic and hydrostatic drag is small, compared to the potential hydrodynamic moment. This is due primarily to low drag, but also because the distance from the slider surface 207 to the pivot center 229 is selected to be relatively small,

at least smaller than the radius of the tip socket 205. The pivot center 229 is also positioned toward the rear of the slider surface 207, which is the preferred location, according to hydrodynamic theory. This tends to tilt the tip 203 forward. The tip 203 is also provided with a straight section 231 on the trailing edg and a slightly larger straight section 233 on the leading edge.

FIGS. 8a and 8b are enlarged diagrams of the tip 203 indicating a typical static-pressure distribution 235 (FIG. 8a) and a typical hydrodynamic pressure distribution 237 (FIG. 8b). FIG. 8a also includes a plurality of arrows that indicate the imposed forces on the tip 203 and FIG. 8b includes a plurality of arrows that indicate the reaction forces on the tip 203. In FIG. 8a, the arrow 219 indicates the rotational direction. The profile of the hydrostaticpressure distribution indicates that the pressure ahead of the tip 203 is greater than the pressure behind the tip 203. Thus, the tip 203 is moving along a section of the cam track 90 where the inlet port is behind and the outlet port ahead of the tip 203. The arrow 239 shows a force, due to pump pressure, acting mainly on the vanetip surface 233 that tends to rotate the tip 203 counterclockwise. The pressure behind the tip 203 results in a force acting mainly on the surface 231 represented by the arrow 241 and tends to rotate the tip 203 clockwise. The arrow 243 represents a hydrostatic pressure from the space in front of the tip 203 that invades the space between the tip 203 and vane 61 and provides a bearing force for the tip 203. The arrow 245 represents the moment due to friction between the vane 61 and tip 203. A frictional drag caused by hydrostatic pressure flow that tends to tilt the tip 203 counterclockwise is represented by the arrow 247. The force due to the hydrostatic pressure distribution 235 that tends to turn the tip 203 clockwise is shown by the arrow 249. FIG. 8b shows an arrow 251 representing the force of the vane 61 tending to hold the vane tip 203 in the socket (205). The arrow 253 represents the reaction moment of the vane due to friction. A drag force due to shear rate in the hydrodynamic film is indicated by the arrow 255. The force due to the hydrodynamic pressures distribution 237 represented by the arrow 257, tends to turn the tip 203 counterclockwise (at this particular set of conditions). However, the position and magnitude of the hydrodynamic pressure force (represented by the arrow 257), varies to balance out the other forces and keep the tip 203 stabilized to travel on, and be supported by, a thin wedge of oil. The friction of the hydrodynamically supported vane-tip 203 is about ten times less than that of boundary (conventional) lubricated vanes. The radius of curvature of the slider surface 207 is selected to be less than the smallest radius encountered in the cam surface 90. The slider surface 207 is selected to be sufficiently large that a hydrodynamic pressure occurring between the tip surface 207 and cam surface 90 is sufficient to counteract other forces acting on the tip from pump pressures and from the vane 61 that supports the tip 203. The straight sections 231 and 233 on the tip aid in balancing the hydrostatic-pressure forces that affect the tip 203. The hydrodynamic action provides a very stiff bearing since the thickness of the hydrodynamic film varies inversely as the square root of the net vane load thus allowing a wide range of forces to act on the tip 203.

FIGS. 9a and 9b show some of the various forces acting on the vane 61. FIG. 9a shows the imposed forces acting on the vane. The hydrostatic force, imposed on the slider surface 207, is represented by the arrow 249 acting on the tip 203. Arrows 238-238 indicate the hydrostatic pressure ahead of the vane 61 acting on the top and side of the vane 61 and also introduced beneath the vane step 209 by the passage 223. The hydrostatic pressure behind the vane 61 is represented by arrows 240240. This pressure also acts on the top, side, and bottom of vane 61. The frictional drag force due to hydrostatic pressure flow is indicated by the arrow 247. Acting through the vane center of gravity 259 are a Coriolis force (arrow 261), a centrifugal force (arrow 263) and an acceleration force (arrow 265); the Coriolis and centrifugal force are actually due to rotor rotation while the acceleration force is an imposed force that depends on the cam track configuration.

The reaction forces are shown in FIG. 9b. The hydrodynamic force (arrow 257) acts through the slider surface 207 on tip 203. Vane socket reaction forces shown by arrows 267 and 269 are imposed on the vane 61 by the rotor 59. A hydrodynamic drag force is represented by the arrow 255.

The vanes 61--61 are responsible for pumping the fluid in the pump 21 and the vane tips 203 provide the support and sealing necessary for the vanes 6161' to operate. The hydrodynamic film beneath the vane surface 207 provides lubrication and support to each vane 61 through the tip 203. Thesteps 209 and 217 and chambers 213 and 215 beneath each vane 61 are means for applying forces to the vane 61 to balance the various other forces that arise due to rotation and the various pressures existing within the pump 21. The construction of the vanes 6161 and their associated tips 203203 are necessary to function properly at the high pressures and high rotational speeds required for the pump 21.

It will be understood, of course, that, while the form of the invention herein shown and described constitutes the preferred embodiment of the invention, it is not intended herein to illustrate all of the possible and equivalent forms or ramifications of the invention. It will also be understood that the words used are words of description rather than of limitation and that various changes, such as changes in shape, relative size, and. arrangement of parts may be substituted without departing from the spirit or scope of the invention herein disclosed.

What is claimed is:

1. In a variable-displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes, comprising:

(a) a plurality of stationary cam track elements;

(b) a plurality of radially moving cam track elements;

and

(c) a plurality of linking elements each having an arcuate surface overlapping and between said stationary and movable cam track elements, said linking elements being constrained to be slidable with respect to said stationary and movable cam track elements, said arcuate surface being nonpivotal during movement of said movable cam track elements and said linking elements moving with said movable cam track elements to vary the size of said pump chamber cam: track.

2. ha variable displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes in accordance with claim 1 wherein the inlet and outlet ports are positioned between the linking elements.

3. In a variable-displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes, comprising:

(a) at least two stationary cam track elements;

(b) at least two radially movable cam track elements, said at least two stationary and at least two movable cam track elements each having a curved section between two straight sections, said straight sections beiing tangent to said curved section, said curved and straight sections being a portion of the entire cam track;

(c) at least one overlapping connecting cam track element disposed between each of said at least two stationary cam track elements and each of said at least two movable cam track elements thereby connecting the stationary and movable cam track elements to complete the pump-chamber cam track, said connnecting cam track element being curved and tangent to the connected straight sections of said movable and stationary cam track elements, each said connecting cam track element being slidable with respect to the stationary and movable cam track elements to vary the amount of overlapping on said straight sections while maintaining tangency, the curved surface of said connecting cam track element being nonpivotal with respect to said stationary and movable cam track elements as said movable cam track elements are moved to vary the displacement of said pump-chamber cam track.

4. In a variable-displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes according to claim 3 wherein said at least two movable cam track elements are interconnected thereby being maintained at equal distances from the central axis of the pump-chamber cam track.

5. In a variable-displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes according to claim 3 wherein said at least two movable earn track elements are pressure balanced by pressure from the pump acting to urge said at least two movable cam track elements toward the central axis of the pump-chamber cam track.

6. In a variable-displacement vane rotary pump, the improvement of providing a pump-chamber cam track for the pump vanes according to claim 3 wherein the inlet and outlet ports of the pump are positioned to pass fluid between each of a movable and stationary cam track element at the position of a connecting cam track element, said ports being positioned along substantially the length of the pump rotor.

7. In a vane rotary pump the improvement of providing a vane tip, comprising:

(a) a slider pivotally mounted in the end of the vane, said slider having a surface contiguous to the inner surface of the pump chamber;

(b) said slider surface being curved with a radius of curvature smaller than the smallest radius of the inner surface of the pump chamber;

(c) the pivotal center of said slider being nearer said curved slider surface than the center point of said curved slider surface and further being nearer the rear of said slider surface; and

(d) said slider surface being selected to be of a size sufiicient to produce a hydrodynamic pressure between said slider surface and the inner surface of 12 said pump chamber, said hydrodynamic pressure being sufiicient to counteract all other forces acting on the tip and to predominate so as to pivotally position said slider surface to form a hydrodynamic load supporting wedge.

8. A vane and rotor construction for a high-speed,

sliding-vane rotary pump, comprising:

(a) a rotor having a plurality of vane sockets;

(b) a vane slidably mounted in each socket of said rotor;

(c) a tip pivotally mounted in the end of said vane, said tip having a curved slider surface opposing the inner surface of the rotor pumping chamber, said curved slider surface having a radius of curvature smaller than the smallest curvature of the inner surface of said rotor pumping chamber thereby inducing a hydrodynamic film between said slider surface and the rotor pumping chamber surface; and

((1) means for imposing forces on said vane to balance forces from the static pressure generated by the pumping action of said vane, said means including a step in said vane and a matching step in said socket to form a first chamber toward the leading edge of said vane and a separate second chamber toward the trailing edge of said vane, said first chamber being larger than said second chamber, said first chamber communicating through the rotor with the pumping chamber in front of said vane as it rotates and said second chamber communicating through said rotor with the pumping chamber behind said vane as it rotates.

References Cited UNITED STATES PATENTS 2,149,337 3/1939 Deming 103136 2,313,075 3/1943 Kendrick et a1. 103120 2,313,246 3/1943 Kendrick et a1. 103120 2,538,193 1/1951 Ferris 10312O 2,538,194 1/1951 Ferris 103-120 3,099,964 8/1963 Eickmann 103-136 3,173,375 3/1965 Eickmann 103136 3,226,013 12/1965 Toyoda et a1 230- 3,285,190 11/1966 Eickmann 103-136 FRED C. MATTERN, JR., Primary Examiner.

WILBUR I. GOODLIN, Assistant Examiner. 

